The present invention relates to reciprocating pumps, and in particular to various types of reciprocating pumps with a linear motor driver and to methods of pumping liquids with such reciprocating pump. Most preferably the pumps of this invention are hermetic reciprocating pumps and the methods of this invention are methods of pumping liquids with such hermetic pumps.
Reciprocating pumps are highly desirable for use in numerous applications, particularly in environments where liquid flow rate is low (e.g., less than 15 gpm) and the required liquid pressure rise is high (e.g., greater than 500 psi). For applications requiring less pressure rise and greater flow rate, single stage centrifugal pumps are favored because of their simplicity, low cost and low maintenance requirements. However, reciprocating pumps have a higher thermodynamic efficiency than centrifugal pumps by as much as 10% to 30%. Although reciprocating pumps are preferred for many applications, they are subject to certain drawbacks and limitations.
For example, traditional reciprocating pumps are commonly driven in a linear direction by a rotating drive mechanism through a slider-crank mechanism or other conventional mechanical mechanism for converting rotary motion to linear motion. These drive systems require multiple bearings, grease or oil lubrication, rotational speed reduction by belts or gears from the driver, flywheels for stabilization of speed, protective safety guards and other mechanical devices, all of which add complexity and cost to the pumps. Moreover, in these traditional constructions the stroke length of the piston is fixed, as is the motion of the piston over time (e.g., generally sinusoidal motion) during each cycle of operation. This results in a peak piston velocity near mid-stroke, which determines the peak Bernoulli effect pressure reduction and kinetic head loss pressure reduction in the fluid that enters the pump on the suction stroke of the piston, thereby effecting the net positive suction head (NPSH) requirement.
Pumps are subject to mechanical damage from insufficient NPSH. In particular, vaporization of liquid at the point of entry into the pump results in vapor bubble formation. Subsequent compression of the vaporized liquid causes violent collapse of the bubbles, resulting in the formation of sonic shock waves that ultimately can damage pump components. Therefore, it is important that the available NPSH of a pump installation be sufficiently above the required NPSH of the pump.
Pump designs requiring a low NPSH allow greater flexibility in installation, often reducing installation costs. In addition, a lower required NPSH assures a greater margin to cavitation and hence greater reliability in operation when inlet operating conditions are off-specification.
The NPSH requirement for reciprocating pumps is dictated by factors tending to reduce the local entry suction pressure, such as liquid line acceleration pressure drop and velocity induced pressure drop (Bernoulli effect and kinetic head losses) in the inlet line and inlet valve. The cylinder and piston size, as well as the inlet valve size and peak piston velocity are critical factors in setting the minimum required NPSH. In particular, larger cylinder, piston and inlet valve size allow a slower pump speed. This results in a lower NPSH requirement. As stated earlier, pump designs requiring a low NPSH allow greater flexibility in installation and also a greater margin to cavitation, both highly desirable attributes.
Adjustment of the speed of traditional reciprocating pumps to reduce the throughput (i. e., flow turndown) is limited largely by the size of the pump flywheel and the size of the electric motor driver. Traditional reciprocating pumps are typically operated at a fixed motor supply power alternating current (AC) frequency and thus a fixed nominal pump speed. Adjustment of the alternating current electrical supply frequency to the motor, such as by the use of a variable frequency drive, to reduce pump speed is typically limited in turndown to 50% of full design pump speed and flow rate. The function of the pump flywheel is to minimize speed fluctuation or ripple during each stroke cycle of the pump. This is accomplished by absorbing and releasing kinetic energy between the pump shaft and the flywheel during each cycle; resulting in a cyclic speed fluctuation of the pump slightly above and below the nominal speed. This is called speed ripple. Speed ripple results in greater and lesser amounts of motor torque at various portions of each pump stroke cycle. This fluctuating torque creates fluctuating motor current draw, which in the extreme can be detrimental to the motor by thermal overheating. The key factor in determining peak motor current draw is the percentage of speed fluctuation. It should be noted that for a given flywheel size and motor size, the speed ripple percentage increases by the square of the ratio of design speed to reduced speed. Additionally, as motor speed decreases, the ability of the motor fan to properly cool the motor decreases as well. These factors combine to create the practical 50% turndown limit. Special measures can be taken to reduce this limit, such as providing a separately powered motor cooling fan, significantly over sizing the pump motor frame or over sizing the pump flywheel. However, these special measures are expensive alternatives. Other means to achieve reduced pump speed, such as variable sheaf diameter belt systems or other mechanical speed ratio adjustment methods, suffer from problems of increased wear, slippage and excessive peak load failures.
When a greater operational flow turndown is required, traditional pumps generally are operated in a recycle mode or in a cyclic on/off mode with a hold up tank. Recycle flow around the pump can be extremely wasteful in pump power and adds cost and complication by requiring a recycle line, a recycle valve, a cooler and means for control. The use of a hold up tank also increases the expense of the system, requires significant excess space and complicates operation and maintenance of the pump system.
A further deficiency associated with traditional reciprocating pumps resides in the need to provide an effective seal between the piston and the pump cylinder. Such a seal typically is provided by piston ring dynamic seals. However, even with the provision of such seals, some leakage is typically encountered, and in many applications represents a nuisance for disposing or recycling of the leaked material.
In traditional reciprocating pumps, piston ring wear is often the primary cause of pump repair maintenance. This results, in part, from sealing the full differential pressure between the pump discharge pressure and the piston backside leakage collection pressure, thereby causing these seals to wear quickly. Specifically, the backside pressure often is equal to or less than the pump inlet pressure, thereby creating a very significant pressure drop across the piston ring seals. This, in turn, increases the resulting piston ring wear rate.
Inlet and outlet valves on a reciprocating pump are typically fluid-activated check valves of specialty design to accommodate the high cyclic rate of the pump while achieving the longest possible operating life. Still, even with the specialty design of these valves, valve failure is often the reason for a pump malfunction. The design speed of the reciprocating pump is based on the required volumetric flow rate and the swept volume of the piston in the pump cylinder. Because a larger swept volume operating at a slower speed requires a larger physical pump size and a higher capital cost, it has been the practice to install a small pump operating at the highest speed permissible, as limited by reciprocating forces, piston ring wear rates and NPSH requirements. Such high speeds, typically in the range of 200 to 600 rpm, place a heavy burden on valve life.
It is desired to have a reciprocating pump that does not have the aforementioned drawbacks of traditional reciprocating pumps, and to actually enhance the positive aspects associated with traditional reciprocating pumps. The reciprocating pumps of the present invention minimize or eliminate traditional reciprocating design drawbacks, including: (1) maintenance of wearing parts, such as valves, piston rings and rod packings; (2) maintenance due to pump cavitation damage in low NPSH applications; (3) leakage of the pumped fluid from the process stream; (4) leakage of the pumped fluid to the pump surroundings; (5) high NPSH requirements for installation design; (6) lubrication contamination of the pumped liquid and pump surroundings; (7) high capital cost; (8) space requirements for installation and (9) hazards associated with exposed moving parts. With the present invention, the aforementioned drawbacks are either minimized or eliminated, while enhancing the positive features of traditional reciprocating pumps, such as high thermodynamic efficiency.
Beneficial aspects of the reciprocating pumps of the present invention that have not heretofore been available include: (1) variable flow from 0% to 100% of design flow rate at full design pressure, with improved efficiency; (2) lower heat leak in cold standby for cryogenic liquid pumping applications; and (3) increased output pressure capability at reduced speed.
Prior art attempts to improve the performance of reciprocating pumps have focused in three (3) areas; namely, modifying the size of traditional slider crank-driven reciprocating pumps, innovative developments in reciprocating cryogenic and/or hermetic pump designs, and converting to linear motor powered reciprocating designs.
With respect to modifying the sizing of traditional slider crank-driven reciprocating pumps, attempts have been made to increase the pump size to provide a swept volume greater than is conventionally considered to be necessary. Employing a bigger pump increases pump costs, but with the benefits of reducing wear-part maintenance by reducing the number of pump cycles required to deliver a predetermined flow, reducing maintenance costs resulting from insufficient NPSH damage, reducing installation costs to meet a high NPSH requirement (e.g., less tank elevation required), and increasing thermodynamic efficiency due to lower speed operation and reduced inlet and outlet valve pressure drop losses.
However, the above stated gains resulting from the use of a larger pump are achieved at the significant expense of: (1) higher pump capital cost; (2) increased fluid leakage from the pumped stream due to the larger piston diameter required to be sealed; (3) increased fluid leakage to the pump surroundings resulting from the larger diameter of the required rod seal; (4) increased general installation costs due to the use of larger-sized parts; (5) increased space requirements due to the use of larger sized parts; (6) increased cost of spare parts; and (7) increased cost of residual maintenance labor due to larger size and handling.
The balancing of the benefits and deficiencies enumerated above has generally resulted in a limitation on the extent of over sizing of reciprocating pumps.
Developments in cryogenic reciprocating pumps have included: (1) employing new dynamic seals, as disclosed in U.S. Pat. No. 4,792,289; (2) modifying the inlet and/or outlet valve designs, as disclosed in U.S. Pat. Nos. 4,792,289; 5,511,955 and 5,575,626; (3) reduced heat leak designs, as disclosed in U.S. Pat. Nos. 4,396,362 and 4,396,354; (4) introducing a second (or multiple) pre-compression chamber(s) for reduced NPSH requirement, as disclosed in U.S. Pat. Nos. 4,239,460; 5,511,955 and 5,575,626; and (5) introducing sub-cooling mechanisms for reducing the NPSH requirement and providing improved volumetric efficiency, as disclosed in U.S. Pat. Nos. 4,396,362; 4,396,354 and 5,511,955. However, none of the above enumerated improvements employ a hermetic design (i.e., no dynamic seals for the pumped liquid to prevent leakage to the ambient surroundings of the pumps).
U.S. Pat. No. 4,365,942 discloses a hermetic cryogenic pump including electrical coils that are maintained superconductive by virtue of the extreme cold temperature of the liquid helium to be pumped. While this design may be unique to the characteristics of liquid helium, it is not widely applicable for use in pumping other fluids.
As noted earlier, other prior art has suggested the use of a linear motor as a driver for a reciprocating pump. Application of this type of driver to a pump has suggested benefits in achieving compact size, reduction of power consumption, reduction of cost, reduction of maintenance and application to situations previously impossible to achieve with traditionally driven pump designs. The use of such linear motor drivers has proven to be applicable to both hermetic and non-hermetic pump designs. Linear motor-powered pumps have been disclosed for use in the down-hole pumping of oil and water, as disclosed in U.S. Pat. Nos. 4,350,478; 4,687,054; 5,179,306; 5,252,043; 5,409,356 and 5,734,209.
U.S. Pat. No. 4,687,054 discloses a wet air gap design that does not employ seals to separate the pumped liquid from the motor's air-gap between the stator and the armature.
U.S. Pat. Nos. 4,350,478; 5,179,306; 5,252,043 and 5,734,209 disclose the use of seals for protecting the motor air-gap from the pumped liquid. Many of the prior art seal designs have the air-gap filled with a lubricating and heat transfer oil. It should be recognized that virtually all of the aforementioned pumps operate fully submerged in the liquid that they pump, and therefore, achieving a hermetic seal to prevent leakage to their ambient surroundings, as desired in the preferred embodiments of the present invention, is a moot point.
Other electric linear motor-driven pumps employing a hermetic design have been disclosed for use in a number of applications, such as for blood pumping (U.S. Pat. No. 4,334,180), large volume, low pressure gas transfer applications (U.S. Pat. No. 4,518,317), a conceptual double-acting pump design (U.S. Pat. No. 4,965,864) and non-hermetic designs employing conventional flat face linear motors (U.S. Pat. No. 5,083,905).
None of the aforementioned prior art teaches a hermetic pump design for intended industrial processes or product delivery applications having all of the benefits of the present invention.
As utilized throughout this application to describe the various embodiments of the invention, the term "swept volume" in reference to the dispensing chamber and/or the reservoir chamber, or in reference to the movement of the piston assembly, refers to the incremental change in volumes of the fluid-receiving regions of the dispensing chamber and reservoir chamber caused by movement of the piston assembly through either a dispensing stroke or a suction stroke. During the dispensing stroke of the piston assembly the volume of the fluid region of the dispensing chamber incrementally decreases by substantially the same amount that the volume of the fluid region of the reservoir chamber increases. During the suction stroke of the piston assembly the volume of the fluid region of the reservoir chamber incrementally decreases by substantially the same amount that the volume of the fluid region of the dispensing chamber increases. The above-discussed incremental decreases and increases in volume of the fluid regions of the dispensing chamber and reservoir chamber are equal to the incremental change in volume of the piston assembly within the dispensing chamber and reservoir chamber as the piston assembly moves through its dispensing stroke and suction stroke, respectively. When the sealing member between the cylinder and piston assembly is fixed against movement to the cylinder, the swept volume equals the traveled distance of the piston assembly moving through the sealing member (in either the dispensing or suction strokes) times (x) the cross-sectional area of that length of the piston assembly which passes through the sealing member.
Reference to "hermetic" or "hermetically sealed" in referring to the various pumps of this invention means pumps that are free of dynamic seals between the pumped fluid and the ambient surroundings of the pump. Dynamic seals are those seals between bodies that move relative to each other with a resulting sliding motion at the sealing point and function to prevent egress of a fluid from a pressurized area to an area of lesser pressure. As stated above, no such dynamic seals are included in hermetic pumps within the scope of this invention between the pumped fluid and the ambient surroundings of the pump.